Double exhaust centrifugal pump

ABSTRACT

The invention relates to a centrifugal pump. The pump ( 10 ) comprises an impeller ( 20 ) and a casing ( 50 ) surrounding the impeller, the impeller having a main passage ( 24 ) that divides into a first passage ( 241 ) and a second passage ( 242 ) with a common feed inlet ( 22 ) parallel to the axis of rotation (A) of the pump, the common feed inlet ( 22 ) being the inlet of the main passage ( 24 ), the first passage ( 241 ) having a first outlet ( 243 ) that is radially oriented, opening out at a first diameter relative to the axis of rotation (A) and suitable for supplying fluid at a first pressure, and the second passage ( 242 ) having a second outlet ( 244 ) that is oriented radially and situated behind the first outlet ( 243 ), opening out at a second diameter that is greater than the first diameter, the second outlet ( 244 ) being suitable for supplying the fluid at a second pressure that is higher than the first pressure, the impeller ( 20 ) co-operating with the casing ( 50 ) to form an axial balancing system having a chamber ( 90 ) formed between the rear face ( 27 ) of the impeller and the portion ( 57 ) of the casing that faces the rear face ( 27 ), the inlet to the chamber ( 90 ) being situated at the level of the second outlet ( 244 ).

The present invention relates to a centrifugal pump.

In the description below, the terms “upstream” and “downstream” are defined relative to the normal flow direction of fluid through the pump and the impeller.

In certain space applications (e.g. rocket engine turbopumps), and in certain industrial applications (e.g. turbopumps in a liquid natural gas (LNG) gasification cycle) make it necessary to have available a first fluid flow at a first rate and at a certain pressure, and another fluid flow at a rate lower than the first rate and at a higher pressure.

In general, two pumps that are fed in series are required for such applications.

FIG. 4 shows a prior art turbopump having two pumps arranged in series, and its operation is summarized below.

A turbopump 100 comprises a rotary shaft 180 suitable for spinning about an axis of rotation A, and a casing 150 surrounding the rotary shaft 180. The casing 150 is prevented from rotation and the rotary shaft 180 is mounted on ball bearings that bear against the casing 150. Around said axis of rotation A there are mounted a primary pump 110, which comprises a primary impeller 112 constrained to rotate with the rotary shaft 180, and a primary exhaust volute 115. The primary exhaust volute 115 is incorporated in the casing 150 and it is stationary.

The fluid penetrates into the turbopump from the front thereof via an annular passage 102 in which an inducer 104 causes the fluid to flow in a direction parallel to the axis A (in the description below, the terms “front” and “rear” are used to define the positions of portions of the pump relative to one another and to a direction parallel to the axis A).

Downstream from this inducer 104 there is a flow-straightening nozzle 106, and downstream from that there is a primary impeller 112. This primary impeller 112 has primary angled passages 104 that move the fluid radially away from the rotary shaft 180 so that the fluid subsequently flows in a direction that is radial relative to the axis A. The fluid is thus entrained into the diffuser 117 situated upstream from the primary exhaust volute 115 and downstream from the primary angled passages 114. The fluid penetrating into the primary volute 115 has been compressed by passing through the primary impeller 112.

Downstream, a fraction of the fluid flowing in the primary volute 115 is taken off and injected into a secondary pump 120 that is situated behind the primary pump 110 along the rotary shaft 180.

The secondary pump 120 has a secondary impeller 122 constrained to rotate with the rotary shaft 180, and a secondary exhaust volute 125. The secondary exhaust volute 125 is incorporated in the casing 150 and it is stationary.

The fluid coming from the primary volute 115 flows into secondary angled passages 124 of the secondary pump 120 that moves the fluid radially away from the rotary shaft 180 so that the fluid then flows in a radial direction relative to the axis A. The fluid is thus entrained into the diffuser 127 situated upstream from the secondary exhaust volute 125 and downstream from the secondary angled passages 124. The section of these secondary angled passages 124 is smaller than the section of the primary angled passages 114. The secondary pump 120 delivers at a rate that is smaller, and at a level of pressure that is higher compared with the fluid flow rate delivered by the primary pump 110.

Nevertheless, having these two pumps leads to a complex arrangement for the turbopump.

Furthermore, the primary impeller 112 includes an axial balancing system that serves to compensate the axial forces (along the axis A) that are generated by the fluid on the rotor. The axial balancing system consists in a chamber 190 situated behind the primary impeller 112, between the impeller and the casing 150. Fluid taken from the outlet of the primary angled passage 114 flows radially through this chamber 190 and leaves the chamber via the portion of the primary impeller 112 that is situated closest to the rotary shaft 180. This flow of fluid through the chamber 190 acts as a cushion opposing the axial forces exerted by the fluid on the rotor. The operation of this axial balancing system requires the rotor to be capable of moving axially relative to the casing 150.

Unfortunately, since the primary impeller 112 delivers pressure that is moderate, the pressure in the chamber 190 is likewise moderate. The take-up capacity of the system depends on the dimensions of the chamber 190 and is a function of the speed of rotation of the primary impeller 112. This speed must not be too high, since otherwise the axial forces exerted by the fluid on the rotor would be too strong to be counterbalanced by the axial balancing system, and that would result in damage to the turbopump.

Such a turbopump is thus of limited effectiveness.

The present invention seeks to remedy those drawbacks.

The present invention seeks to provide a pump of size, weight, and cost that are reduced, and of performance that is increased.

This object is achieved by the fact that the pump includes an impeller and a casing surrounding the impeller, the impeller having a main passage that divides into a first passage and a second passage with a common feed inlet parallel to the axis of rotation of the pump, the common feed inlet being the inlet of the main passage, the first passage having a first outlet that is radially oriented, opening out at a first diameter relative to the axis of rotation and suitable for supplying fluid at a first pressure, and the second passage having a second outlet that is oriented radially and situated behind the first outlet, opening out at a second diameter that is greater than the first diameter, the second outlet being suitable for supplying the fluid at a second pressure that is higher than the first pressure, the impeller co-operating with the casing to form an axial balancing system having a chamber formed between the rear face of the impeller and the portion of the casing that faces the rear face, the inlet to the chamber being situated at the level of the second outlet.

By means of these provisions, the turbopump presents size and weight that are smaller since the primary pump and the secondary pump are replaced by a single pump having two exhausts. There is thus no longer any need to take off fluid from the outlet of the primary pump in order to feed the secondary pump. Thus, the configuration of the turbopump is simplified and fabrication costs are reduced. Furthermore, since the diameter of the impeller is greater, the performance of the axial balancing system associated with this two-exhaust pump is improved, as explained below.

Advantageously, the first passage and the second passage are separated by a shroud having at least a portion that extends between the first outlet and the second outlet.

The invention can be better understood and its advantages appear better in the light of the following detailed description of an embodiment given by way of non-limiting example. The description refers to the accompanying drawings, in which:

FIG. 1 is a longitudinal section view of a turbopump including a centrifugal pump of the invention;

FIG. 2 is a longitudinal section view of a centrifugal pump of the invention;

FIG. 3 is a longitudinal section of another embodiment of a centrifugal pump of the invention; and

FIG. 4 is a longitudinal section view of a turbopump of the prior art.

FIG. 1 shows a turbopump 1 having a rotary shaft 80 suitable for spinning about an axis of rotation A, and a pump 10 mounted on said rotary shaft 80. The pump 10 has an impeller 20 that is constrained to rotate with the rotary shaft 80, and a casing 50 that is prevented from rotation, and that surrounds the rotary shaft 80. The rotary shaft 80 is mounted on ball bearings that bear against the casing 50.

The fluid penetrates into the turbopump 1 via the front thereof through an annular duct 2 leading to an inducer 4 that is followed by a flow-straightening nozzle 6, with the impeller 20 being situated downstream therefrom. The inducer 4 and the nozzle 6 form parts of the pump 10.

The impeller 20 has a main passage 24 with a feed inlet 22 into which the fluid from the inducer 4 penetrates. The feed inlet 22 is oriented along the axis A. Downstream from this feed inlet 22, the main passage 24 forms a bend going away from the axis of rotation A, where it splits into a first passage 241 and a second passage 242. The first passage 241 leads to a first outlet 243 that is thus oriented substantially perpendicularly to the axis of rotation A.

The separation between the first passage 241 and the second passage 242 is provided by a shroud 40.

As shown in FIG. 2, the shroud 40 is situated behind the first passage 241 (i.e. the first passage is closer to the front of the turbopump), and therefore forms the rear wall of the first passage 241 as far as the first outlet 243.

The shroud 40 is situated in front of the second passage 242, and thus forms the front wall of the second passage 242. The second passage 242 extends radially beyond the level of the first outlet 243 so as to terminate in a second outlet 244 that is further from the axis A than is the first outlet 243. The external portion 48 of the shroud 40 is the portion of the shroud 40 that extends between the first outlet 243 and the second outlet 244.

As shown in FIG. 2, the shroud 40 has an internal portion that extends the external portion 48 upstream over the major portion of the angled region of the passage 24. The first passage 241 and the second passage 242 are therefore angled, and of cross-section (perpendicularly to the flow along the passage) that converges slightly between the upstream end 45 and respectively the first outlet 243 and the second outlet 244.

The upstream end 45 of this internal portion of the shroud 40 is situated downstream from the feed inlet 22. For example, this upstream end 45 is situated closer to the feed inlet 22 than to the first outlet 243.

Ideally, the thickness of the shroud 40 is as thin as possible so as to optimize the separation of the fluid in the passage 24.

Alternatively, as shown in FIG. 3, the shroud 40 does not extend towards the axis A beyond its external portion 48. Under such circumstances, the upstream end 45 of the shroud 40 is situated at the same distance from the axis A as the first outlet 243.

The shroud 40 may also extend upstream within the main passage 24 over a distance that is intermediate between the configuration shown in FIG. 2 (the shroud extends almost as far as the inlet 22 of the main passage 24) and the configuration shown in FIG. 3 (the shroud does not extend upstream beyond the first outlet 243).

The casing 50 forms a volute having two exhausts, including a first exhaust volute 51. The first outlet 243 faces a first orifice 513 of the first exhaust volute 51, so that the fluid is driven by the impeller 20 towards the first exhaust volute 51 through the first orifice 513.

The casing 50 also includes a second exhaust volute 52. The second outlet 244 is situated facing a second orifice 524 forming part of the second exhaust volute 52, such that the fluid is entrained by the impeller 20 to the second exhaust volute 52 through the second orifice 524.

The second passage 242 is of cross-section that is smaller than that of the first passage 241.

Thus, the first passage 241 supplies a flow at a large rate and at a moderate pressure, while the second passage 242 supplies a flow at a smaller rate (since its cross-section is smaller) and at a higher pressure (since the second outlet 244 is radially further away from the axis A than is the first outlet 243).

There is therefore no longer any need to take fluid from the outlet of the exhaust volute, as is done in the exhaust volute 115 of the prior art (FIG. 4). Consequently, the section of the first exhaust volute 51 may be smaller than the section of the exhaust volute 115.

The second exhaust volute 52 is adjacent to the first exhaust volute 51, behind it and further away from the axis of rotation A. This configuration serves to minimize the volume occupied by the casing 50.

The external portion 48 of the shroud 40 of the impeller 20 faces towards the front of the pump. Facing this external portion 48, there is an intermediate wall 58 that forms a portion of the casing 51 and that separates the diffuser 511 of the first exhaust volute 50 from the diffuser 522 of the second exhaust volute 52. Since the intermediate wall 58 forms a portion of the casing 50, it is stationary.

When the pump is in operation, the external portion 48 of the shroud 40 is driven in rotation about the axis of rotation A and it therefore moves relative to the intermediate wall 58. Undesirable fluid leakage thus occurs at the interface between this external portion 48 and the intermediate wall 58 due to the pressure difference between the first outlet 243 and the second outlet 244 of the impeller.

It is desirable to minimize this leakage flow.

Given the relative movement between this external portion 48 and this intermediate wall 58, sealing of this interface is preferably provided by a dynamic sealing gasket. For example, the gasket may be a labyrinth gasket.

As shown in FIGS. 2 and 3, the front wall 29 of the main passage 24 includes a dynamic gasket at the inlet 22 of the passage 24, at the interface between said front wall and the casing 50. This gasket is designed to limit leaks, and thus to limit the flow that is recirculated between the first outlet 243 of the first passage 241 and the inlet 22 and that needs to be recompressed. By way of example, this gasket may be a labyrinth gasket.

Alternatively, and in particular in applications where the performance of the turbopump is less crucial, the front wall 29 of the passage 241 and/or the shroud 40 and its external portion 48 may be omitted.

Behind the rear face of the impeller 20 there is a single axial balancing system that is used for compensating the axial forces (along the axis A) that are generated by the fluid on the rotor when the pump is in operation.

The rear wall of the second passage 242 is formed by the impeller 20, and the rear face 27 of the impeller 20 faces a portion 57 of the casing 50 over its entire area. The space between this rear face 27 and this portion 57 forms a chamber 90.

Fluid is taken from the second passage 242 level with the second outlet 244, and it enters into this chamber 90 via an annular passage 93 of variable axial clearance, leaving it through an outlet orifice 96 situated level with the portion of the impeller 20 that is situated closest to the rotary shaft 80. The assembly comprising the annular passage 93 and the chamber 90 constitutes the axial balancing system.

The operation of such an annular passage 93 of variable axial clearance is summarized briefly below.

The fluid penetrates via the annular passage 93 into the chamber 90. The chamber 90 fills up since the pressure difference between the annular passage 93 and the outlet orifice 96 ensures that the fluid flows into the chamber 90. This fluid exerts pressure towards the front that tends to move the impeller 20 forwards. The shape of the annular passage 93 is such that this forward movement of the impeller tends to close the passage. The pressure in the chamber 90 therefore decreases, thereby allowing the impeller 20 to move rearwards.

It can thus be understood that the axial balancing system enables the impeller 20 to be maintained in its axial position about an equilibrium point. The axial balancing system thus has a function of regulating the axial position of the impeller 20 (and thus of the rotary shaft 80). This system therefore has the advantage of being an active system, in contrast to a passive system in which the compensation force against the impeller is independent of the axial position of the rotor and the rotary shaft.

In FIGS. 1 to 3, the annular passage 93 includes a step, in a known configuration.

It can thus be understood that the axial balancing system enables the impeller 20 to be maintained in its axial position about an equilibrium point.

Given that the radial height of the impeller 20 is greater than the height of an impeller of the prior art, the area of the chamber 90 against which the pressure acts is greater. Furthermore, the pressure difference between the annular passage 93 and the outlet orifice 96 is greater. The pressure in the chamber 90 can thus vary to a greater extent, and the axial balancing system can thus accommodate greater axial forces. This considerable increase in the capacity of the axial balancing system enables the speed of rotation of the turbopump to be increased, while conserving sufficient take-up capacity.

The resulting advantages comprise size, weight, and thus cost for the turbopump that are reduced, together with turbopump efficiency that is greater.

Furthermore, there is no need for the impeller 20 to be coupled to the rotary shaft 80 (as in the prior art) in order to accommodate the operation of the axial balancing system. Given that the pump 10 has only one impeller, this forms a single block together with the rotary shaft 80 (as shown in FIG. 1) and it is the rotary shaft 80 and the impeller 20 that move together axially in order to accommodate the operation of the axial balancing system.

Fabrication of the pump of the invention is thus simplified, since it includes fewer parts.

In FIGS. 1 to 3, the cross-section of the second passage 242 is shown as being smaller than the section of the first passage 241. The opposite configuration is also possible, providing the pressure at the outlet from the second passage 242 remains greater than the pressure at the outlet from the first passage 241. 

1. A pump, characterized in that it comprises an impeller and a casing surrounding the impeller, said impeller having a main passage that divides into a first passage and a second passage with a common feed inlet parallel to the axis of rotation of said pump, said common feed inlet being the inlet of said main passage, the first passage having a first outlet that is radially oriented, opening out at a first diameter relative to the axis of rotation and suitable for supplying fluid at a first pressure, and the second passage having a second outlet that is oriented radially and situated behind said first outlet, opening out at a second diameter that is greater than said first diameter, said second outlet being suitable for supplying the fluid at a second pressure that is higher than said first pressure, said impeller co-operating with said casing to form an axial balancing system having a chamber formed between the rear face of said impeller and the portion of said casing that faces said rear face, the inlet to said chamber being situated at the level of the second outlet.
 2. The pump according to claim 1, characterized in that said first passage and said second passage are separated by a shroud having at least a portion that extends between said first outlet and said second outlet.
 3. The pump according to claim 2, characterized in that said shroud is extended upstream into said main passage towards said common feed inlet.
 4. The pump according to claim 2, characterized in that the external portion of said shroud that extends between said first outlet and said second outlet is provided with a dynamic sealing gasket at its interface with an intermediate wall of said casing.
 5. The pump according to claim 1, characterized in that the interface between the front wall of said main passage and said casing is provided with a dynamic sealing gasket.
 6. The pump according to claim 1, characterized in that it further includes an inducer and a flow-straightening nozzle upstream from said common feed inlet.
 7. The pump according to claim 3, characterized in that: the external portion of said shroud that extends between said first outlet and said second outlet is provided with a dynamic sealing gasket at its interface with an intermediate wall of said casing; the interface between the front wall of said main passage and said casing is provided with a dynamic sealing gasket; and it further includes an inducer and a flow-straightening nozzle upstream from said common feed inlet. 